The shaft or rotor of high-speed rotary machinery is often supported by bearings so that the shaft may rotate at very high speeds such as up to about 150,000 revolutions per minute (RPM), while retaining its alignment. There are a number of phenomena which act on the shaft to throw it out of alignment and the effects of these phenomena are generally more severe at higher rotational speeds.
One such phenomenon is often termed synchronous whirl and is caused essentially by centrifugal forces acting on a mass-unbalanced shaft. The shaft is generally mass-unbalanced because typically the geometric and inertial axes of the shaft are not identical due to such factors as machining tolerances and material imperfections. Further unbalance may result from repeated assembly and disassembly of multiple elements, from normal wear and from shaft deflection during operation. At certain rotational speeds, the shaft tends to rotate on its inertial rather than its geometric axis causing the shaft to orbit or whirl within the bearing housing. When the synchronous whirl excites a natural frequency in the bearing system, the system's vibrational amplitudes can become excessive. The system's natural resonant frequency is generally referred to as its critical speed. Since a long shaft supported, for example, by two axially spaced journal bearings can have many lateral modes of vibration, the system can have many critical speeds. Consequently, there may be further cases of synchronous whirl resonance at the second, third or higher critical speeds depending upon the frequencies of the rigid and bending modes of vibrations of the rotor-bearing system.
Another such phenomenon is commonly referred to as half-frequency, half-speed or self-excited whirl. As the shaft speed approaches a speed approximately equal to twice the critical speed, the shaft experiences a harmonic vibration or whirl which is superimposed on the synchronous shaft whirl. This vibration rapidly increases in amplitude and is often catastrophic to the bearing.
Other phenomena which may cause shaft instability, especially in overhung or straddle-mounted rotary machinery such as turbocharges, cryogenic expanders, compressors, expander-driven compressors and the like, are aerodynamic induced excitations which may originate, for example, from pressure variations around the circumference of impellers and seals.
Still another source of rotating shaft instability may be forces stemming from material hysteresis, rubbing between rotating and stationary parts and other such activity common to rotary machinery.
To ensure that the effects of phenomena such as those described do not lead to bearing and perhaps even machine failure, the bearing system must be able to support the shaft while also effectively counteracting the tendency of the shaft to vibrate or to move radially.
One way for the bearing system to counteract the shaft vibrations is by a dual cushion system comprised of springs and viscous fluid. As the shaft moves radially it exerts a force in the direction of movement on the bearing across their common length which force may or may not be equal across this common length. The bearing system resists and cushions this force by exerting a force on a spring mechanism in contact with it and a stationary surface, and on a viscous fluid between the bearing system and the stationary surface. The different responses to the exerted force by the spring and the viscous fluid tend to dampen the vibration. In such a system, shaft support is provided by the bearing, flexible support is provided by the spring mechanism and damping is provided by the viscous fluid.
One bearing system known to the art is described in U.S. Pat. No. 4,097,094--Gardner. In this Gardner system a journal bearing is resiliently supported in a rigid housing by arcuate spring elements. Viscous damping is provided by a fluid-filled gap between the bearing and the housing. This gap has an axial length less than that of the bearing surface.
The Gardner system has several disadvantages. One disadvantage is that because the amount of viscous damping is directly related to the axial length of the fluid-filled gap, the amount of viscous damping which can be provided is not totally independent of the bearing surface. This is because in the Gardner system the fluid-filled gap cannot have an axial length greater than the bearing surface and thus to achieve a greater amount of viscous damping than that provided by any given fluid-filled gap, one would have to increase the axial length of the bearing surface. This is undesirable because the increased bearing surface introduces a mechanical penalty to the rotary machinery and also because dynamic response problems may arise due to the altered bearing surface and shaft surface relationship.
Another disadvantage of the Gardner system is the limitation on the total amount of damping attainable. As is known, the amount of damping can be increased by decreasing the width of the fluid-filled gap. However, the width of the gap cannot be reduced to less than about 0.001 inch because of machine tolerances. Moreover, thermal distortion of parts, dirt entrained within the viscous fluid and conical excursions of the bearing housing make it difficult to maintain even this size gap. Thus, the Gardner system is constrained in the amount of damping attainable without increasing the gap axial length and unavoidably the bearing surface length.
As previously discussed, at certain rotary speeds the shaft excites a natural frequency in the bearing system which can lead to catastrophic results. One way to avoid such results is to design the bearing system so that its natural frequencies do not coincide with the desired operating rotational speeds of the machinery. However, one is constrained in the design of bearing systems by the need for the bearing system to provide effective shaft support, flexible support and viscous damping.
One way to overcome this constraint is to provide a bearing system wherein these three functions can be adjusted independently of each other. Therefore a change in bearing system design to, for example, avoid the natural frequency problem which may effect one of the bearing system functions would not effect the other two functions. In such a system one could design the bearing system or change the design so as to have a beneficial effect on one function without encountering the possibility of an unavoidable detrimental effect on one of the other functions.
It is therefore an object of this invention to provide an improved bearing system for rotary machinery.
It is another object of this invention to provide an improved bearing system for rotary machinery wherein the shaft support, flexible support and viscous damping functions are provided independently of one another such that each function can be adjusted individually without affecting either of the other two functions.